Guide flanges in roller bearings, such as e.g. tapered roller bearings, can be embodied either straight or spherical. Straight flanges are mainly used for roller bearings of smaller diameters and thus also smaller flange widths. In this case, a flange width is often too small to produce a defined profile on an available flange surface. Spherical flanges, i.e. flanges having a constant curvature, find application mostly in roller bearings having larger diameters and thus also having larger flange widths. A spherical flange is characterized in that a radius, which defines a flange shape facing towards a roller raceway, has its origin substantially on an axis of rotation of the rollers (roller rotational axis), wherein small deviations are allowed due to alignment errors.
For a more detailed description of roller bearings having spherical flanges, FIG. 1 shows, in a schematic representation, a longitudinal section of a roller bearing 10, which is formed in an exemplary manner as a tapered roller bearing. The roller bearing 10 includes a bearing inner ring 11, a bearing outer ring 12, and a plurality of rollers 13, which can roll on races or raceways 14, 15 formed by the inner sides of the bearing rings 11, 12. In the case of a tapered roller bearing, tapered rollers are the rollers corresponding to the rollers or rolling elements 13.
The tapered rollers 13 can roll on an inner raceway 14, which is formed in the bearing inner ring 11, and on an outer raceway 15, which is formed in the bearing outer ring 12. In a tapered roller bearing, the raceways 14, 15 are formed as conical outer surfaces. In the longitudinal section shown in FIG. 1 of the tapered roller bearing 10, the raceways 14, 15 define, in an imaginary extension, an inner line 16 and an outer line 17, which meet on an axis of rotation 18 of the roller bearing 10 ideally at a center of rotation 19.
During operation of the bearing 10, each (tapered) roller 13 rotates around its own roller axis 20, wherein an imaginary extension of the roller axis 20 ideally also intersects with the center of rotation 19. For the tapered rollers 13, a rolling condition on the raceways 14, 15 is realized by a relative position of the inner line 16, outer line 17, bearing axis of rotation 18, and roller axis 20, which all intersect in the center of rotation 19, so that during a relative rotation of bearing inner ring 11 and bearing outer ring 12, the tapered rollers 13 roll on the raceways 14, 15 substantially without slippage, and an amount of friction related thereto is minimized.
When supporting axial forces, in order to also optimize the friction that occurs in the axial direction i.e. in the direction of the bearing rotational axis 18, the rollers 13 used in roller bearings can have a first curvature, on their end side 21, identified by a first radius R21, so that a part of a ball surface results as an end side surface of the rollers 13, which, as is indicated in the enlargement of FIG. 1, is in contact at a contact point 22 with a straight- or spherically-embodied flange 23, for example of the bearing inner ring 11. Away from the region of a possible contact point 22, the curved end side can also have a second curvature defined by a second radius R23, which curvature is less than the first curvature of the spherical end side surface 21 of the rollers 13. In this region the rollers can also be flat. Curvature is generally understood to mean the change in direction per traversed length of an infinitesimally short curved piece. A circle having the radius r thus has the same, constant curvature 1/r everywhere; its direction changes everywhere equally strongly. With all other curves the curvature can vary from curve point to curve point, or along a path on the one three-dimensional surface. The inverse of the curvature is referred to as the radius of curvature. This is the radius of that circle (circle of curvature) which represents the best approximation of the observed curve in the vicinity of the contact point.
In roller bearings, such as for example cylindrical roller-, barrel roller-, or ball-bearings, which are designed with straight or flat flanges, in comparison to spherically embodied flanges the roller-flange contact has a higher surface pressure (Hertzian pressure) between the roller end side 21 and the flange surface facing towards this. Here the Hertzian pressure is understood to be the greatest pressure that prevails in the middle of the contact surface of two elastic bodies. If, such as with roller bearings having straight flanges, two elastic bodies (curved roller end side and straight or flat flange) are pressed against each other, then in the ideal case they touch only in a punctiform manner. However, in the real case, a flattening and thus a contact surface arises at the contact point 22 due to the elasticity. A characteristic pressure distribution (surface pressure) arises on the contact surface in both bodies, wherein the pressure is always highest in the middle. If, as here, a ball outer surface and a flat flange surface touch, a touch- or contact-ellipse results. Due to the comparatively high surface pressure, with roller bearings having straight flanges, a relatively poor lubricant film formation generally results at higher effective forces. In addition, in comparison to spherical flanges, straight or flat flanges lead to smaller contact ellipses between the roller end side and the flange surface facing this, for which reason an overlapping of the contact ellipse with the flange edges can result only at extreme loads. Likewise, with straight- or flat-embodied flanges, there is a low sensitivity of the contact point 22 to alignment errors, so that a defined contact point 22 between roller 13 and flange is possible. While on the one hand a greater skewing of the roller 13 is made possible with flat-embodied flanges, on the other hand a relatively poor guiding of the rollers results during operation.
Tapered roller bearings in the large bearing field can, as shown with reference to FIG. 1, be embodied with spherical flanges 23, which compared to straight or flat flanges results in a lower surface pressure between the roller end side 21 and the flange surface or abutment surface facing towards the roller 13 as a consequence. In addition, spherically-designed flanges 23 lead, compared to straight flanges, to larger contact ellipses between the roller end side 21 and the opposing flange surface, so that overlappings of the contact ellipse with the flange edges and thus edge stresses can frequently result. In general, with spherically-designed flanges 23 there is a higher sensitivity of the contact point 22 to alignment errors than is the case with flat- or straight-embodied flanges. Although on the one hand spherical flanges have a lower skewing of the roller 13 as a consequence, on the other hand due to the narrow osculation between the roller end side 21 and the flange surface facing towards roller 13, the roller 13 can be guided better during operation. A defined contact point 22 between the roller 13 and the flange 23 is also theoretically possible with spherical flanges by a different choice of the radii of curvature (and/or their origins) of roller end surface 21 and spherical flange surface.
However, one of the main disadvantages of a spherical flange is the resulting sensitivity of the contact point 22 between the roller end side 21 and the flange 23 to alignment error. Deviations in the raceway angle, roller angle, flange radius, as well as roller end side radius have a decisive influence thereon.